Defect and Diffusion Forum
ISSN: 1662-9507, Vol. 401, pp 117-130
© 2020 Trans Tech Publications Ltd, Switzerland
Submitted: 2019-05-24
Revised: 2019-10-02
Accepted: 2020-04-15
Online: 2020-05-28
Turbulent Heat Transfer Characteristics of a W-Baffled Channel Flow Heat Transfer Aspect
Y. Menni1,a*, A.J. Chamkha2,b, O.D. Makinde3,c
Unit of Research on Materials and Renewable Energies, Department of Physics, Faculty of
Sciences, Abou Bekr Belkaid University, P.O. Box 119-13000-Tlemcen Algeria
1
Mechanical Engineering Department, Prince Sultan Endowment for Energy and Environment,
Prince Mohammad Bin Fahd University, Al-Khobar 31952, Saudi Arabia
2
Faculty of Military Science, Stellenbosch University, Private Bag X2, Saldanha 7395, South Africa
3
a*
menniyounes.cfd@gmail.com, bachamkha@pmu.edu.sa, cmakinded@gmail.com
Keywords: Forced convection, Nusselt number, thermal enhancement, w-baffle, heat transfer
aspect.
Abstract. In this work, the thermal behavior of a turbulent forced-convection flow of air in a
rectangular cross section channel with attached W-shaped obstacles is investigated. The continuity,
momentum and energy equations employed to control the heat and velocity in the computational
domain. The turbulence model of k-ε is employed to simulate the turbulence effects. The finite
volume method with SIMPLE algorithm is employed as the solution method. The results are
reported temperature, local and average Nusselt numbers, and mean velocity contours. The subject
is relevant and important for industrial applications.
Introduction
The arrangement of obstacles, such as baffles, fins and ribs, within channels are among the
effective methods used by many researchers and investigators in their numerical and experimental
studies. Muszyński and Kozieł [1] carried out two-dimensional numerical investigations of the fluid
flow and heat transfer for the laminar flow of the louvered fin-plate heat exchanger, designed to
work as an air-source heat pump evaporator. The simulations were performed for different
geometries with varying louver pitch, louver angle and different louver blade number. The
maximum heat transfer improvement interpreted in terms of the maximum efficiency was obtained
for the louver angle of 16° and the louver pitch of 1.35 mm. Park et al. [2] systematically presents
the results of heat transfer and friction factor data measured in five short rectangular channels with
turbulence promoters. They investigated the combined effects of the channel aspect ratio, rib angleof-attack, and flow Reynolds number on heat transfer and pressure drop in rectangular channels
with two opposite ribbed walls. Their experimental results were also compared with literature
values. Experimental and numerical studies were conducted by Wong et al. [3] to investigate the
dynamic and thermal behavior of a turbulent airflow in a horizontal air-cooled rectangular duct,
with inclined square-sectioned cross-turbulators mounted on its bottom surface. Effects of varying
the angle formed by the cross-turbulators between 30° and 120° on the convective heat transfer and
pressure drop were studied. An optimum range of angles formed by the cross-ribs corresponding to
a maximum enhancement of forced convection was observed. According to the experimental and
numerical results obtained, its value would be between 60° and 70°. Nasiruddin and Kamran
Siddiqui [4] indicated that the convective heat transfer in a heat exchanger tube may be enhanced by
placing a baffle inside the tube. The investigators considered a comparative study between three
different baffle orientations. The first case examined a vertical baffle. The second case investigated
a baffle inclined towards the upstream end, and the third one considered a baffle inclined towards
the downstream end. The results suggested that a baffle inclined towards the downstream side with
a smaller inclination angle (15° in their study) is a better choice as it enhances the heat transfer by a
similar magnitude with a minimal pressure loss. Hwang and Liou [5] investigated heat transfer and
loss friction in a rectangular channel with symmetrically mounted solid and fully perforated ribs on
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Computational Analysis of Heat Transfer in Fluids and Solids II
parallel broad walls (rib open-area ratio: β = 50%; rib pitch-to-height ratio: Pi/H = 5-20; rib heightto-channel hydraulic diameter ratio: H/De = 0.081 and 0.162; rib-to-channel height ratio: H/2B =
0.13 and 0.26; rib height-to-channel hydraulic diameter ratio: H/De = 0.081 and 0.162; Reynolds
number: Re = 10,000-50,000). The results indicated that the perforated ribs had the advantages of
eliminating the hotspots and providing a superior heat transfer performance.
Karwa and Maheshwari [6] investigated fully (open area ratio of 46.8%) and half (open area
ratio of 26%) perforated baffles covering Re values ranging from 2,700 to 11,150. The study
showed an enhancement of 79-169% in Nusselt number over the smooth duct for the fully
perforated baffles and 133-274% for the half perforated baffles while the friction factor for the fully
perforated baffles are 2.98-8.02 times of that for the smooth duct and is 4.42-17.5 times for the half
perforated baffles. In general, the half perforated baffles are thermo-hydraulically better to the fully
perforated baffles at the same pitch. Of all the configurations studied, the half perforated baffles at a
relative roughness pitch of 7.2 give the greatest performance advantage of 51.6-75% over a smooth
duct at equal pumping power. Sahel et al. [7] presented a new baffle design to eliminate the
formation of lower heat transfer areas (LHTAs), particularly in the downstream regions of solidtype baffles. This design concerns a perforated baffle having a row of four holes placed at three
different positions. These positions are characterized by a ratio called the pore axis ratio (PAR)
which is taken equal to 0.190, 0.425 or 0.660. In their study, the baffle perforated with PAR =
0.190 was found to be as the best design, which reduces significantly the LHTAs, since it ensures
an increase in the thermal transfer rate from 2% to 65 %, compared with simple baffle. However,
the pressure loss may decrease until 12 times compared with the simple baffle. Khan et al. [8]
described experimental investigation of heat transfer with turbulent flow in a rectangular channel
with inclined solid and perforated baffles combined with rib turbulators. Combining ribs with
perforated and inclined baffles yielded an increase in average Nusselt number, albeit with a pressure
drop penalty. In situations where rate of heat transfer is critical to the performance of a device,
combining ribs with baffle is a viable solution. Se Kyung et al. [9] have done a numerical study on
the heat transfer and frictional characteristics of airflow inside a rectangular channel fitted with
different types of inclined baffles (type I: solid baffle; type II: 3 hole baffle; type III: 6 hole baffle;
and type IV: 9 hole baffle). The numerical results of the flow field showed that the flow patterns
around the different baffle configurations are entirely different and these significantly affect the
local heat transfer characteristics. The heat transfer and friction factor characteristics are
significantly affected by the perforation density of the baffle plate. It was found that the heat
transfer enhancement of baffle type II has the best values. An experimental investigation was
carried out by Ko and Anand [10] to measure the heat transfer coefficients and pressure loss in a
uniformly heated rectangular channel with wall mounted staggered porous baffles. The experiments
were conducted in Reynolds number range of 20,000-50,000. The use of porous baffles resulted in
heat transfer enhancement as high as 300% compared to heat transfer in straight channel with no
baffles. The experimental result analysis showed that the heat transfer enhancement per unit
increase in pumping power was less than one for the range of parameters studied. Guerroudj and
Kahalerras [11] simulated the influence of porous block shape on the laminar mixed convective heat
transfer and airflow characteristics inside a two-dimensional parallel plate channel when the
buoyant and forced flow effects are simultaneously present. The influence of the buoyancy force
intensity, the porous blocks shape going from the rectangular shape to the triangular shape, their
height, the porous medium permeability, the Reynolds number and the thermal conductivity ratio
was analyzed. The results revealed essentially, that the shape of the blocks can alter substantially
the flow and heat transfer characteristics. A computer code was developed by Kamali and Binesh
[12] to study the turbulent heat transfer and friction in a square duct with various-shaped ribs
mounted on one wall. The simulations were performed for four rib shapes, i.e., square, triangular,
trapezoidal with decreasing height in the flow direction, and trapezoidal with increasing height in
the flow direction. The results showed that features of the inter-rib distribution of the heat transfer
coefficient are strongly affected by the rib shape and trapezoidal ribs with decreasing height in the
flow direction provide higher heat transfer enhancement and pressure drop than other shapes.
Defect and Diffusion Forum Vol.401
119
Sripattanapipat and Promvonge [13] simulated the laminar periodic flow and heat transfer in a two
dimensional horizontal channel with isothermal walls and with staggered diamond-shaped baffles.
They reported that the diamond shape of the baffle with different tip angles (5 to 35°) may enhance
the heat transfer from 200 to 680% for Reynolds number ranging from 100 to 600. However, this
intensification is associated with enlarged friction loss ranging from 20 to 220 times above the
smooth channel. Saini and Saini [14] conducted an experimental prediction on the turbulent flow
and convective heat transfer characteristics in a rectangular air channel with arc-shaped elements
attached to the underside of a heated plate. The effects of various dimensionless parameters such as
arc angle and height on the Nusselt number and friction factor were studied for Reynolds numbers
ranging from 2,000 to 17,000. Their results suggested that a significant heat transfer coefficient
enhancement in a solar air channel can be achieved by introducing arc-shaped ribs into the flow.
The maximum enhancement in Nusselt number was obtained as 3.80 times corresponding to the
relative arc angle of 0.3333 at relative roughness height of 0.0422, with the minimum pressure loss.
Stehlik et al. [15] compared heat transfer and friction loss correction factors of an optimized
segmental baffle heat exchanger to those of a helical baffle heat exchanger. In their studies, the
correction factors for helical baffles were examined as a function of baffle inclination angle to gain
an understanding of the underlying transport phenomena as well as to characterize the baffle for
design purpose.
An improved structure of ladder-type fold baffle was proposed by Wen et al. [16] in order to
block the triangular leakage zones in original heat exchangers with helical baffles. They
numerically showed that the distribution of shell-side velocity and temperature in improved heat
exchanger are more uniform and axial short circuit flow is eliminated. The fluid flow and heat
transfer characteristics of the improved heat exchanger and the original heat exchanger were also
experimentally studied. They showed that the shell-side heat transfer coefficient and overall heat
transfer coefficient are improved by 22.3-32.6% and 18.1-22.5%, respectively. A numerical
investigation for fully developed turbulent flow in a square duct fitted with 45° in-line V-baffle
pairs mounted on both upper and lower walls was conducted by Fawaz et al. [17] in order to
examine the changes in flow structure and thermal performance, using air as the working fluid at Re
ranging from 5,000 to 25,000. Effect of various baffle blockage ratios (BR = 0.2, 0.4 and 0.6) and
baffles pitch ratios (PR = 0.5,1 and 1.5) on flow behavior and heat transfer were investigated. They
found that the TEF of the V-baffle pointing upstream at BR = 0.2 is higher than that at larger BR
and the TEF of this same baffle at PR = 0.5 is higher than that at higher PR, at the lowest Re value.
Sriromreun et al. [18] reported experimental and numerical investigations of the heat transfer and
flow friction characteristics for a solar air heater channel with in-phase and out-phase Z-shaped
baffles in the turbulent regime from Re = 4,400 to 20,400. The Z-baffles inclined to 45° relative to
the main flow direction are characterized at three baffle-to channel-height ratios (e/H = 0.1, 0.2 and
0.3) and baffle pitch ratios (P/H = 1.5, 2 and 3). The effects of e/H and P/H ratios were more
significant for the in-phase Z-baffle than for the out-phase Z-baffle. In addition, numerical [19] and
experimental [20-25] works on air baffled channels were conducted to analyze the heat transfer
coefficient and pressure loss. In those studies, different structural parameters of the model and
various operating parameters were used.
In this paper, a computational thermal analysis of the turbulent forced-convection fluid stream
behavior in the presence of two staggered, transverse, solid-type, double V-form baffles (or Wshaped obstacles) is conducted in order to improve the heat transfer phenomenon within thermal
devices. The thermal aspect is shown for flow Reynolds numbers based on the hydraulic diameter
of the channel ranging from 12,000 to 32,000. To perform a detailed analysis of the aerodynamic
and thermal fields within this model of W-baffled channel, the finite volume approach, by means of
Commercial CFD software FLUENT, for a steady-state, incompressible, and two-dimensional flow
is used, and the SIMPLE-algorithm is implemented for all calculations. The analysis results are
presented in terms of mean velocity, fields and profiles of temperature, local and average numbers
of normalized Nusselt, and thermal enhancement factor for various axial and transverse channel
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Computational Analysis of Heat Transfer in Fluids and Solids II
stations. Small attack angle value (α = 45°) and high Reynolds number (Re = 32,000) values lead to
the best functioning regime in a channel containing W-shaped obstacles.
Problem Definition
The particularity of this computational thermal analysis is the forced-convection heat transfer
aspect. The main aim of this paper is to examine the turbulent flow field and convective heat
transfer characteristics of a constant property Newtonian fluid inside a two-dimensional horizontal
channel of rectangular form, containing two transverse, staggered, solid-type, double V-form (or Wshaped) baffles, where a constant surface temperature is applied on all solid boundaries of the duct,
Fig. 1.
Figure 1. Geometry under examination (dimensions in m).
This channel section due to of their specific geometry, accelerate the flow disturbance and
improve the heat exchange between the working fluid and the hot walls. The flow Reynolds number
is varied between 12,000 and 32,000.
The physical models for fluid flow in the rectangular cross section channel provided by double
V-baffles were developed under the following assumptions:
1- Steady two-dimensional heat transfer and fluid flow;
2- Flow is turbulent and incompressible;
3- Physical properties of air fluid (Cp, μ, λf, ρ) and solid (λs) are constants;
4- Fluid is viscous Newtonian;
5- Temperatures applied to the lower and upper surfaces of the channel (Tw) are considered
constants;
6- Velocity profile at the inlet of the channel is uniform;
7- Radiation heat transfer is not considered;
8- Thicknesses of the bottom and top wall surfaces of the channel are neglected; and
9- Standard k-ε turbulence model proposed by Launder and Spalding [26], by means of Commercial
CFD software FLUENT is applied in this thermal analysis to simulate the fluid flow behavior.
The governing flow equations, i.e., continuity, momentum and energy equations, used to
simulate the incompressible steady fluid flow and heat transfer in the whole domain treated are
given as [4, 29, 30]
(1)
∇V = 0 ,
2
(2)
ρ V .∇V = −∇P + µ f ∇ V ,
(
)
ρc (V .∇T ) = k ∇ T .
p
f
2
(3)
The standard k-ε model is defined by two transport equations, one for the turbulent kinetic energy, k
and the other for its dissipation rate ε, as given below [26]



∂
(4)
(ρkui ) = ∂  µ + µt  ∂k  + Gk − ρε
σ k  ∂x j 
∂xi
∂x j 
and
Defect and Diffusion Forum Vol.401

µt
 µ +
σε

The turbulent viscosity, μt is modeled as follows:
∂
(ρεui ) = ∂
∂xi
∂x j
 ∂ε 
ε
ε2

 + C1ε Gk − C2ε ρ
k
k
 ∂x j 
121
(5)
k2
(6)
ε
In these equations, Gk represents the generation of turbulence kinetic energy due to the mean
velocity gradients. The related constant parameters are [26]
(7)
C1ε = 1.44, C2ε = 1.92, Cµ = 0.09, σ k = 1.0, σ ε =1.3
µt = ρCµ
The Reynolds number is defined as
ρ U Dh
µ
The local Nusselt number, Nux which can be written as
h D
Nu x = x h
Re =
λf
(8)
(9)
The average Nusselt number, Nu can be obtained by
1
(10)
Nu = ∫ Nu x ∂x
L
where hx is the local convective heat transfer coefficient. In the entrance region of the test section,
1- A uniform velocity profile is applied, u = Uin and v = 0;
2- At the bottom (y = -H/2) and top (y = H/2) channel walls, the non-slip and impermeability
boundary conditions are implemented over the channel wall as well as the W-baffle surfaces, that is
u = v = 0;
3- In the channel outlet (x = L) it is prescribed the atmospheric pressure, P = Patm;
4- The inlet temperature of air Tin is considered to be uniform at 300 K;
5- A constant temperature of Tw = 375 K is applied on the entire walls of the channel as the thermal
boundary condition.
CFD Solution
The two-dimensional, incompressible Navier–Stokes equations and the turbulence model
equations are discretized using the finite volume method, details of which can be found in Patankar
[27]. For the momentum equations, the pressure and velocities are linked together based on the
SIMPLE-algorithm [27]. Considering the characteristics of the flow, the Quick-scheme [28] was
applied to the interpolations, while a Second-order upwind scheme [27] was used for the pressure
terms. In order to verify the accuracy of numerical results obtained in this numerical study with the
computer code Fluent, a validation of our analysis was made by comparing with the results of
Demartini et al. [25] that are available in the literature. These authors studied a similar problem for
the circulation of air through a rectangular channel but with simple baffles. Under the same
conditions, we conducted a comparison in terms of the axial velocity profiles for a Reynolds
number equal to Re = 8.73×104 at axial location x = 0.159 m from the entrance, shown in Fig. 2.
The comparison of results by our numerical method and results of Demartini et al. [25] for steadystate flow conditions shows very good agreement as illustrated in the figure.
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Computational Analysis of Heat Transfer in Fluids and Solids II
Figure 2. Numerical validation of dimensionless axial velocity profiles at station x = 0.159
m for Re = 8.73×104.
Results and Discussion
Fig. 3 shows the contour plots of mean velocity fields for different Re numbers, i.e., Re =
12,000, 17,000, 22,000, 27,000, and 32,000. It is clearly noted that the fluid velocity values are
almost negligible near the two W-obstacles, particularly in the downstream areas this is caused by
the presence of the recirculation zones. Far from these regions, the current lines become parallel,
which leads to the progressive development of the flow. It is also worth noticing that the velocity
increases in the region extending from the end of each W-obstacle to the wall of the channel. This
rise in velocity is caused by the presence of the W-obstacles and also by the presence of recycling;
hence, an abrupt change in the direction of the flow comes out. One can also notice that the largest
velocity values are found near the top of the channel. The flow starts accelerating just after the
second W-obstacle. The velocity of the flow is also influenced by the Re number value. If the Re
number is increased, then the flow accelerates in the vicinity of the W-obstacle faces, and this
causes the convective heat transfer rate to rise. The analysis of the isotherms presented in Fig. 4 for
various values of Reynolds number, shows that the fluid temperature significantly increases with
the presence of the double V-baffles compared to that of the situation with no baffles. In the
downstream region of the two double V-baffles, recirculation cells with relatively high temperatures
are observed. In the space between the tip of each double V-baffle and the walls of the channel, the
temperature is decreased. The total temperature profiles are shown in Fig. 5 (a) to (d) for different
transverse sections of the channel. The numerical results show that the temperature of the air in the
recirculation zone is substantially high compared to that obtained in the same region without
baffles. This observation is confirmed by Nasiruddin and Kamran Siddiqui [4]. We also note that
the hottest areas are, mostly, located near the solid boundaries of the test channel (lower and upper
walls of the channel, and double V-baffle surfaces). It is also found that the temperature value at the
heated wall level decreases with increasing the flow velocity. On the other hand, the results analysis
allowed associating to the fluid temperature elevations, the effect of double V-baffles and fins. At
the output of each free segment between the tip of each double V-baffle and the channel walls, the
air in flow does not encounter any metal obstacles; its velocity decreases due to the sudden
enlargement, and the lack of W-baffles constitutes a supplementary factor of local attenuation of the
turbulence in these areas. However, the air temperature increases as soon as the fluid once again
finds in contact with the double V-baffles, and this is repeated in an analogous manner from one
cell to another. What was also noticed, the fluid temperature is inversely related with the increase in
the Reynolds number. The graphical representations of the temperature variation as a function of
Reynolds number (Re = 12,000; 22,000; and 32,000) in the transverse section situated between both
the W-baffles at x = 0.255 m, is shown in Fig. 6. These results certify that the heat exchange
between the heat transfer fluid and the heated walls in the channel with staggered 45° double V-
Defect and Diffusion Forum Vol.401
123
baffles is more important with low flow Reynolds numbers. It is clear that for high Reynolds
numbers, the fluid temperature significantly decreases i.e., there exists an inverse proportionality
between increasing flow Reynolds number and the total temperature in each channel cross section.
Additionally, according to analysis of our numerical results on the velocity profiles [see Fig. 6] and
the total temperature profiles for different sections of the channel [see Figs. 5 and 6), it is found that
the fluid temperature is related to the flow velocity.
(a)
(b)
(c)
(d)
(e)
Figure 3. Mean velocity fields for various values of flow rate: (a) Re = 12,000,
(b) Re = 17,000, (c) Re = 22,000, (d) Re = 27,000, and (e) Re = 32,000.
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Computational Analysis of Heat Transfer in Fluids and Solids II
(a)
(b)
(c)
(d)
(e)
Figure 4. Temperature fields for various values of flow rate: (a) Re = 12,000,
(b) Re = 17,000, (c) Re = 22,000, (d) Re = 27,000, and (e) Re = 32,000.
(a)
Defect and Diffusion Forum Vol.401
125
(b)
(c)
(d)
Figure 5. Profiles of fluid temperature (a) upstream of the first W-baffle at x = 0.159 m and
x = 0.179 m, (b) between the first and the second W-baffles at x = 0.255 m and x = 0.285 m,
(c) before the second W-baffle at x = 0.315 m and x = 0.335 m, and (d) near the channel outlet
at x = 0.525 m, Re = 12,000.
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Computational Analysis of Heat Transfer in Fluids and Solids II
Figure 6. Variation of fluid temperature profiles with Reynolds number for x = 0.255 m.
Figure 7. Distribution of normalized local Nusselt number along the bottom and top walls of the
channel for Re = 12,000.
The heat transfer rate, characterized by the normalized local Nusselt number, is then
determined and shown along the lower and upper walls of the channel with Re = 12,000 in Fig. 7.
These profiles present in all cases (bottom or top walls) a minimum and a maximum of the Nusselt
number. It is found that the heat transfer rate minimums are observed at the level of base of these
double V-baffles and that the Nusselt number increases along the baffle and reaches its maximum
on its upper face. The effect of the Reynolds number on the normalized local Nusselt number
evolution is seen in Fig. 8 (a) and (b) for both bottom and top walls of the channel, respectively.
The results show that the heat transfer rate is increased with the increase of flow Reynolds number,
because when the Re increases, the turbulence increases and the recirculation region become
stronger and consequently the heat dissipation increases. If we think in terms of no-dimensional
average Nusselt number, as shown in Fig. 9, there is increasing almost linearly Nusselt number as a
function of Reynolds number. There is a linear increment between average heat transfer ratio and
flow Reynolds number value. It may be generated by the augmentation of the flow acceleration
causing by increasing the flow velocity. It is then found that the double V-baffles play an effective
factor to dissipate the heat from the solid walls and the temperature of the flow increases in the
regions occupied by the double V-baffles and between the W-deflectors. This enhances the heat
transfer.
Defect and Diffusion Forum Vol.401
127
(a)
(b)
Figure 8. Distribution of normalized local Nusselt number along the (a) bottom and (b) top
walls of the channel for different flow Reynolds numbers.
Figure 9. Variation of normalized average Nusselt number with Reynolds number.
Fig. 10 shows the variations of the thermal enhancement factor (TEF) as a function of the
Reynolds number at the lower and upper channel walls. In the figure, the TEF value tends to
increase with augmenting the Re number for both channel surfaces. The upper wall containing the
first 45° W-shaped baffle provides the highest TEF value while the lower wall fitted with the
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Computational Analysis of Heat Transfer in Fluids and Solids II
second 45° W-shaped baffle yields the lowest one. It is found that the TEF values vary between
0.853-1.037; and 1,277-1,644 for lower and upper walls of the channel, respectively, depending on
the Re values. In the case of bottom surface, the TEF was found to be much larger than unity; its
maximum value was around 1,644 for Re = 32,000. This value of TEF is decreased by 36,952 % in
the case of the bottom channel wall at same Re number.
Figure 10. Evaluation of thermal-aerodynamic performance for various Reynolds numbers.
Conclusion
The most important conclusions that can be drawn from this study are as follows:
• The velocity of the flow is influenced by the Re number value. If the Re number is
increased, then the flow accelerates in the vicinity of the W-obstacle faces, and this causes
the convective heat transfer rate to rise;
• The temperature of the air in the recirculation zone is substantially high compared to that
obtained in the same region without baffles;
• In the space between the tip of each double V-baffle and the walls of the channel, the
temperature gradient is increased;
• At the output of each free segment between the tip of each double V-baffle and the channel
walls, the air in flow does not encounter any metal obstacles; its velocity decreases due to
the sudden enlargement, and the lack of W-baffles constitutes a supplementary factor of
local attenuation of the turbulence in these areas.
• The fluid temperature is inversely related with the increase in the Reynolds number.
• The heat transfer rate minimums are observed at the level of base of these double V-baffles
and that the Nusselt number increases along the baffle and reaches its maximum on its upper
face
• The double V-baffles play an effective factor to dissipate the heat from the solid walls and
the temperature of the flow increases in the regions occupied by the double V-baffles and
between the W-deflectors. This enhances the heat transfer.
• The TEF values vary between 0.853-1.037; and 1,277-1,644 for lower and upper walls of
the channel, respectively, depending on the Re values.
Defect and Diffusion Forum Vol.401
129
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